Braking system for a light rail vehicle

ABSTRACT

A brake system for a light rail vehicle having both dynamic and friction braking systems is disclosed. Further, this light rail vehicle has both motored and nonmotored axles. The system has several individual and redundant systems which may act independently on each truck. The system uses a dynamic brake means which converts the propulsion motors to brake motors as the prime braking means for the motored axles. The system includes a fluid operated friction brake system with a proportional control means to supplement the dynamic braking with proportional friction braking as needed to achieve the total braking force required. The system includes an air spring and variable load means for each truck to regulate the maximum amount of fluid pressure available for the friction brake means to prevent over application of the friction brakes and skidding of the wheels. The system further includes parking brake and emergency brake means with independent actuators to effect actuation of the friction brakes.

ied States Patent 1191 [111 3,924,902

Eagle 5] Dec. 9, 1975 BRAKING SYSTEM FOR A LIGHT RAIL VEHICLE PrimaryExaminer-Duane A. Reger [75] Inventor: Thomas Engle, Cape Vincent, NY.33 52 Agent or Flrm polflock Phllpltt & Vande [73] Assignee: GeneralSignal Corporation,

Rochester, NY. 57 ABSTRACT 1 [22] Filed: June 17, 1974 A brake systemfor a light rail vehicle having both dynamic and friction brakingsystems is disclosed. Fur- [211 Appl' 48044l ther, this light railvehicle has both motored and non- Related U.S. Application Data motoredaxles. The system has several individual and [63] Continuation-impart ofSer. No. 393,529, Aug. 31, redundant systems which y act independentlyon 1973, Pat. No, 3,345,991, each truck. The system uses a dynamic brakemeans which converts the propulsion motors to brake motors [52] U.S. Cl.303/21 A; 303/22 A as the prim king mean for he mo or axl s. Th [51]Int. Cl. B60T 8/04 y m incl es a flu op ra e friction br ke ystem [58]Field of Search 188/170, 181 A; 303/6 C, with a proportional controlmeans to supplement the 303/3, 13, 15, 21 A, 22 R, 22 A, 21 P dynamicbraking with proportional friction braking as needed to achieve thetotal braking force required.

[56] References Cited The system includes an air spring and variableload UNITED STATES PATENTS means for each truck to regulate the maximum3 443 842 5/1969 Pier 303/21 AX amount of fluid pressure available forthe friction 3:501:203 3/1970 PaneIIIIIIIIII I: 303/22 A x bTake meansto i l over apphcatmn of the 3,599,761 s/1971 Schultz et al. 188/170brakes and Sklddmg of the Wheels- The System 3,658,390 4/1972 Chouings303/21 A further includes Parking brake and emergency brake 3,754,7958/1973 Vo Lewis et a1 303 22 A UX means with independent actuators toeffect actuation 3,791,702 2/1974 Burckhardt et al 303/21 A x of thefriction brakes. 3,799,297 3/1974 Ryburn et a1. 188/170 6C 7D 3,833,2719/1974 Partitt et al 303/21 A 1 guns 9 E/ 5 ms t n v guns men/1cm INPUTSmay f 27 .111) SIGNAL BRAKE 20 CENTER TRUCK m To 33% 00mm anus comm commFRICTION ouu 19 l/ 20 l i |6b L- 30 30b AMPLIFIER AMPLIFIER LONTROLVLVCONTROL VLV US. Patent Dec. 9, 1975 Sheet 2 of4 3,924,902

.3. Patent Dec. 9, 1975 Sheet 3 of4 3,924,902

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BRAKING SYSTEM FOR A LIGHT RAIL VEHICLE CROSS-REFERENCE TO RELATEDAPPLICATION This application is a continuation-inpart of Ser. No.393,529 filed Aug. 31, 1973, now Pat. No. 3,845,99l issued Nov. 5, 1974.

BACKGROUND OF THE INVENTION The present invention relates to a brakingsystem for use on a light rail vehicle, and more particularly, a masstransit of rapid transit car. Relatively complex braking systems havebeen used for passenger cars in the past, but the new rapid transit carsare substantially lighter than either standard passenger cars or freightcars. This results in a substantial change in the load to empty weightratio of the vehicle. The load to empty weight ratio for older stylepassenger cars and freight cars does not change appreciably even whenthe vehicle is fully loaded. However, light weight rapid transit carsmay carry a load which is equal in weight to that of the car bodyitself. This change in the load to empty weight ratio, together with thefrequent and rapid stops of the rapid transit cars require a relativelymore complex and sophisticated braking system. V

The system of the present invention is particularly intended for use inelectrified vehicles wherein dynamic or regenerative braking isutilized. It is common practice to use the propulsion means of thevehicle as the braking means by exciting the traction motor fields andallowing the motor means to act as a generator when dynamic braking isrequired. The present invention is intended to provide a system whichwill combine the advantages of dynamic braking with the precision offriction braking through a proportional control system.

The wide disparity in empty weight to load ratios also createsadditional problems in friction braking. The amount of force required toeffect a given speed reduc tion for a heavily loaded vehicle which causethe wheels of a lightly loaded vehicle to slide on the rails. This notonly reduces the effective braking, but also creates flat spots on thewheel which require repair or replacement of the wheel and a reductionof the inservice time of the car. Accordingly, it is necessary to useproportional braking systems that will limit the maximum braking effortof the friction brakes in accordance with the weight of the vehicle.

In recent years, rapid transit or mass transit cars have used airsprings or air spring suspension systems rather than coil springs tosupport the vehicle. The air spring systems provide superior ridecharacteristics, and will also maintain the height of the car body afixed distance above the tracks and the loading platform regardless ofthe load in the cars. The present invention presumes the use of airsprings for the proportional braking systern.

The present invention is also concerned with a parking brake for thevehicle. The engineering of the parking brake presents two problems forthe designer. It is preferable to have a quick acting parking brakecontrol separate from the friction brake control system to provide aredundancy for brake operation. Similarly, it is desired to have a handbrake control which can be applied without the necessity for externalenergy from any source other than the operator himself.

It is desirable to provide slip or skid detectors on mass transit orrapid transit vehicles. Since the stops are relatively frequent, and thetrain often encounters a variety of rail surfaces in a short distance,it is necessary to provide means for breaking the skid andreestablishing the brake force after the skid has been terminated.

It is also desirable to provide a completely redundant emergency brakesystem with a separate control system and separate actuators to effectapplication of the friction brakes in an emergency situation.

SUMMARY OF THE INVENTION The braking system of the: present invention isintended for use on a light rail vehicle of the mass transit or rapidtransit variety. The braking system has several individual brakingsystems which may act independently and redundantly on each axle. Morespecifically, the brake system of the present invention utilizes anindependent dynamic braking system and an independent pneumaticallycontrolled proportional friction brake system. The present inventionprovides proportional control means for adding the proper amount offriction braking force to the force established by the dynamic brakesystem. The system uses the dynamic brake as a primary brake on motoredaxles, while using the friction brake on all nonmotored axles so as toutilize the available adhesion on all wheels during stopping to providethe shortest stop distance without sliding wheels. On the motored axlesthe friction brake is energized only when the dynamic brake is incapableof producing the selected braking effort, and then only to the extentnecessary to satisfy the deficiency.

The control system for the present invention also includes a hand brakeor parking brake system with completely independent means for applyingthe friction brake. The parking brake utilizes a hydraulicallyrestrained spring motor which actuates the braking pads when the parkingbrakes are applied. The parking brakes may be applied by merely ventingor releasing the hydraulic motor which restrains the spring motor.

The emergency brake system for the present invention consists of meansfor applying the friction brake in the full service mode and also forapplying track brakes. Thus, if the system is operating properly, thehydraulically restrained spring motor will never come into play.However, the hydraulic restraining pressure for the spring motor isconnected to the service hydraulic brake cylinder line during emergencybraking, so that should the service brake cylinder pressure fail, thefact of this failure would vent oil pressure from the hydraulicallyrestrained spring motor and cause the spring applied brake to actuate.Use of the spring applied brake as a back-up during emergency brakeapplications assures that there can be no failure or combination offailures, which will deprive the car of an emergency brake. At the sametime use -of the normal full service brake as the emergency frictionbrake in the absence of failure assures that the braking will beconsistent with vehicle weight; that is, will be load limited. Thespring applied brake is not load limited, and thus would cause slippingwheels on a lightly loaded car with a possible consequence of increasedstopping distance; thus, its use is relegated to a back-up brake to comeon only in the event of a hydraulic or pneumatic failure.

In addition to the three redundant systems disclosed above, there aretwo additional systems which are capable of independent intervention inthe application of the friction brakes. The first is the variable loadsystem which regulates the amount of fluid pressure available to thefriction brake actuating mechanism. A variable load valve and air springare provided for each truck. The system uses the air spring pressure toproduce an output pressure of a minimum threshold value, or of a valueproportional to the air spring pressure above the threshold. Thisproportional application of pneumatic pressure matches the applicationof the friction brake to the weight of the car.

It is also an object of the present invention to provide a completelyindependent wheel slip subsystem to release the brakes on any slidingtruck, and then reapply the brakes when the wheel slip has beenterminated. This wheel slip subsystem operates by comparing the speedsof all axles on the car and if the speed of one of the axles exceedsthat of the others by more than a predetermined amount, it energizes adump valve on the skidding truck to release the brakes.

The present invention is intended to provide a brake system for lightrail vehicles, wherein a dynamic or regenerative brake means generates avariable electric control signal upon application of the dynamic brakemeans. The system also includes a fluid operated friction brake meansfor the vehicle with an actuating mechanism which is responsive tovariations in fluid pressure to actuate the friction brake means. Thepresent invention also includes a fluid pressure responsive controlmeans responsive to a variable electric control signal to providevariations in the fluid pressure applied to the friction brake means.The fluid pressure control means includes an electropneumatic pressuretransducer which responds to the variations in the electrical controlmeans to vary the fluid pressure supplied to the friction brake means.It is also an object of the present invention to provide an actuator forthe friction brake comprising first and second pressure responsivehydraulic motors, with the first hydraulic motor responsive to positivefluid pressure variations to actuate the friction brake means. A springmotor is also provided and is responsive to reductions in hydraulicpressure supplied to the second hydraulic motor to actuate the brakingmeans.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a functional block diagramof the interrelationship between the various components of the brakesystem of the present invention.

FIG. 2 is a cross sectional and schematic view of the air pressurecontrol manifold, and the fluid pressure proportional control valve.

FIG. 3 is a cross sectional view of the variable load valve. I

FIG. 4 is a cross sectional view of the pneumatic to hydraulic converterand dump chamber.

FIG. 5 is a cross sectional view of the dump valve and a portion of thedump chamber.

FIG. 6 is an isometric and cross sectional view of the manifold and dumppassageway.

FIG. 7 is a diagrammatic and partially cross sectional view of theemergency and handbrake control system, the hydraulic actuator and thedisc brake.

DESCRIPTION OF THE PREFERRED EMBODIMENT To aid in understanding thepresent invention, the entire brake control system, and the interactionof the various related components will be described with respect tofunctional block diagram FIG. 1. Each of the various subsystems willthen be described in detail in connection with the description of thestructural components of that subsystem.

FIG. 1 illustrates the interrelationship between the electricalsubsystem and the pneumatic subsystem. In addition, it illustrates theinterrelationship between the three independent braking systems, and thetwo independent intervening systems. The three independent brake systemscomprise the dynamic braking system, the pneumatically controlledfriction brake system, and the independent emergency and parking brakecontrol system. The independent intervening systems include the variableload system and the wheel slip detection and release system.

The braking system of the present invention is intended for use in masstransit or rapid transit, and will customarily involve only a limitednumber of cars. This is contrasted with conventional brake systems whichare intended for use in freight or passengers trains involving or 200cars. In the light rail vehicle for which the preferred emobidment isintended, propulsion units and dynamic brakes 28 are included for atleast the two end trucks of the vehicles. In addition, each of thedynamic brakes provides a signal feedback to a brake command control 29which responds directly to changes in current in the control line system11. The design of the command control and the dynamic brake may takemany different forms, and the illustration in FIG. 1 is intended to berepresentative of a wide variety of forms. It is understood that thecomplete dynamic brake also includes a master controller (not shown)which is normally mounted in the lead unit of the train which permitsthe operator to switch the propulsion control circuits of the tractionmotors between motoring and dynamic braking configurations.

The present invention also includes a pneumatically controlled frictionbrake system utilizing friction brakes 32 and 33 for the front truck, 34and 35 for the center truck, and 36 and 37 for the rear truck. Thepresent invention uses a pneumatic control system and pneumatic tohydraulic convertors 38, 39 and 40 for converting the pneumatic controlsignals to hydraulic pressure. These convertors amplify the pneumaticsignals and facilitate the use of hydraulic actuators 32-37. Thehydraulic portion of the braking system also includes slack adjusters41, 42 and 43 to provide the increased precision required with a fastresponse braking system.

Each of the pneumatically controlled friction brake motors is responsiveto variations in pressure control. Each of the convertors 38-40 includesa pressure responsive fluid motor and a hydraulic motor. The fluid motoris responsive to positive variations in fluid pressure and actuates thehydraulic motor. The fluid pressure for the pneumatically controlledfriction brake system is independently varied by two separate systems.The initial incoming air pressure is varied by the control valve 45which provides a brake pressure output that is proportional to areduction in electric current input below a preset level. This signalcurrent is provided by a brake command control unit 29 and is reduced inresponse to an operator call for braking and increased by either anoperator call for brake release, or in the case of powered (dynamicallybraked) axles by an increase in dynamic brake feedback current. In thebrake system of the present invention, the dynamic brake is used as theprimary brake on the powered axles and the friction brake is energizedonly when the dynamic brake is incapable of producing the necessarybraking effort, and then only to the extent necessary to satisfy thedeficiency. The control valve 45 is in electropneumatic pressuretransducer which responds to control signals from the brake commandcontrol 29, to produce an output pressure which regulates aproportioning bypass valve. The control valve 45 provides a brakingsystem which is able to increase and decrease the friction brakingeffort as required during a brake application to produce exactly thetotal braking effort called for by the operator through control line 11,regardless of the irregularities in the dynamic brake re sponseattributable to speed effects or malfunctions.

The maximum fluid pressure that can be achieved is also independentlyvaried by the variable load valves 53a, 53b, and 53c. The operation ofthe variable load circuitry will be subsequently described.

The control valve 45 controls the fluid pressure available to themotored end trucks of the vehicle, while control valve 46 provides fluidpressure for controlling the application of brakes 34, 35 on thenonmotored center truck. This brake is applied independently of thedynamic braking applied to the end trucks. The center truck provides ofthe braking effort, while the end trucks provide 40% at each truck. Thesystem generates air for the pneumatic control system by means of aircompressor 47 which feeds the main storage reservoir 48 and the brakingreserve reservoir 49. This reservoir is protected against loss of airclue to pipe breakage or compressor malfunction by check valve 50. It isalso sufficiently large to provide 15 full service brake applicationsand releases with the air compressor 47 completely inoperative. Supplyreservoir 49 supplies the pressure for the pneumatic control systemwhich in turn actuates the friction brake system. The main reservoir 48also supplies air pressure to the air springs 51a, 5th, and 510 locatedbetween the car trucks and the car body. Check valve 50 isolates thepneumatic control system for the variable load system, to prevent theloss of control pressure in the event of a rapture in one of the airsprings, or a malfunction of one of the variable load components.

in normal operation, the variable load valves 53a, b, c regulate themaximum amount of pressure available to the pneumatic to hydraulicconvertors 3840. This is necessary since during a full brakeapplication, a predetermined amount of braking force is applied to thewheels of the vehicle through the friction braking systern to bring thevehicle to a stop as quickly and as safely as possible. lt is important,however, that the braking force not be excessive since this would causethe wheels to slide on the tracks. The sliding and skidding results inthe formation of flat spots on the wheels and a lengthening of thedistance required to bring the car to a halt. Since the braking forcerequired during a full brake application is proportional to the totalweight of the car, including its load, it is necessary to provide meansfor measuring the load and regulating the brake cylinder pressure duringa full service application. If this were not done, full braking forcerequired for a fully loaded car would cause the wheels of an empty carto slide, or conversely the full braking force required for an empty carwould be insufficient to quickly and safely stop a fully loaded car.

The car body is normally isolated from the trucks by means of airsprings which serve several functions. The primary purpose is to providea spring with ride characteristics which are superior to those of a coilspring. Additionally, the air spring mechanism will maintain the heightof the car body a fixed distance above the tracks and the loadingplatform regardless of the load in the car. This is accomplished byvarying the pressure of the air spring mechanism in accordance with thechange in the load carried by the car body. The variable air pressurepresent in air spring 51a, 51b, and 51c may then be used to provide aproportional indication of the load carried by the vehicle itself. Thisvariable pressure is used to pilot the variable load valves 53a, 53b,and 53c which will vary the amount of pressure supplied to the pneumatict-o hyraulic convertors 3840.

The brake system of the present invention also provides hydraulicactuators for the friction brakes which have first and second hydraulicmotors arranged in tandem. The first hydraulic motor is responsive topositive variations in hydraulic pressure to actuate the disc brakes.The second hydraulic motor is responsive to reductions in hydraulicpressure to release a spring motor means mounted in the brake cylinder.The spring motor means effects a mechanical application of the brakepads to the brake disc.

The first hydraulic motor means is actuated by the control systempreviously discussed. The electric or electronic control signals aretranslated to pneumatic control signals which in turn actuate thepneumatic to hydraulic convertors 3840. These convertors providepositive variations in the hydraulic pressure supplied to the firstmotor means mounted in the brake cylinders 3237. This system isproportional and applies a graduated brake application depending uponspeed, load, and dynamic braking available.

As will be subsequently explained, the emergency backup and parkingbrake control system comprises a completely separate subsystem thatoperates independently of the service brake control system. Both theemergency backup brake system and the parking brake systems use thespring motor means and the second hydraulic motor means. Each is appliedby de-energizing the emergency magnet valve which in the absence ofservice brake cylinder pressure will reduce the hydraulic pressure inthe second hydraulic motor means. This pressure is released by saidemergency magnet valves 301-303 attached to each of the vehicle trucks.These valves are normally closed, zero leakage magnet valves whichestablish communication between the second fluid motor and the hydraulicsupply line. If no pressure is present on the supply line, the trappedpressure is quickly dissipated through the convertors 38-40. If aservice brake application is in progress, the pressure is equalized toprevent an over application of brake pressure and subsequent skidding ofthe vehicle.

A sub-system is also provided for recharging the second hydraulic motorduring each service brake applica tion. This insures adequate pressurein the second hydraulic motor means and prevents inadvertentapplications of the emergency brakes through leakage.

Exhaust valves 301-303 are actuated by an electric signal impressed oncontrol line 304 by either the emergency brake control 14 or the parkingbrake control 10. In addition, means are also provided for releasing thebrakes manually when desired.

The braking system of the present invention is also provided with awheel slip or skid detection system comprising the wheel slip detectorlogic circuitry 27, and wheel slip dump valves 59. and 61. The logiccircuitry 27 will detect differences in speed between the axles on thevarious trucks to signal for a reduction 7 in with the dynamic brakesystem, or the friction brake system. The requirements for fail safeoperation are met through having normally de-energized relays to providethe circuit connections between the wheel slip logic 27 and the dynamicbrakes 28 and 29, and normally de-energized dump valves 59-61. The logiccontrol circuitry further provides electrical timing means for the dumpvalves which prevents their remaining open for more than a set periodafter encrgization. Thus. when a skid is detected, the dump valves areenergized, and release the pneumatic pressure in the pneumatic tohydraulic convertor for the slipping or skidding truck. The convertor onthe'affected truck or trucks is dumped for the length of time necessaryto correct the slip, but not longer than the preset period.

The wheel slip detectors utilize signal inputs from four magnetic pickupsensors. The logic circuitry 27 amplifies, shapes and compares the lowlevel signal pulses from the wheel slip detectors. While the detectionsystem will operate to produce a spin or slide indication when any axleis rotating at a speed different from that of any other axle by morethan a small amount, there is also a provision to detect situations inwhich all wheels slide simultaneously at synchronous speeds. This isdone by detecting the change in rate of angular acceleration. The systemmeasures the acceleration of one axle and when triggered will cause thedetector to operate and send a dump signal to a signal dump valve. Thedump valve operation will correct the skid for only one truck, but assoon as that truck begins to correct its slide, the synchronism isbroken and the other detectors will actuate their respective dumpvalves.

The system also includes a service brake cut-out valve 63a, forcompletely de-energizing the friction brake system. The cut-out valve isused for towing and or completing the mission in the event ofmalfunction of one of the braking systems. Manual cut-out is alsoprovided to each truck on a per truck basis by cut-out valves 63b, 63cand 63d.

PROPORTIONAL CONTROL SYSTEM The proportional control system of thepresent invention has three independent subsystems. The first employs aproportional control valve to mix the friction braking with the dynamicbraking for the end trucks. The second applies a friction only brakeeffort to the center truck. The third proportional system includes airsprings and variable load valves for each of the vehicle trucks.

In operation, the vehicle P wire control 9 supplies a variable signalfrom O to 10 volts to the brake command control center. The centeradjusts a number of factors including the total vehicle load weight, thejerk limit for the train, and the amount of dynamic brake feedbackreceived from the dyanmic brake means. The control center then suppliesa graduated signal to the proportional control valve 45. The amplifier30 boosts the incoming control signal to a O to volt range for use bythe proportional control valve. The proportional control valve receivesa predetermined input pressure from conduits l6 and 16a of approximately100 psi. The proportional control valve produces an output pressurewhich increases as the voltage from amplifier decreases from theprescribed level. This proportional control valve uses many of thecomponents described and illustrated in my US. Pat. Nos. 3,528,709,Electric Current to Pneumatic Pressure Transducer,

8 3,536,360 and 3.536.361 entitled Blending Scheme for CurrentResponsive Railway Brake, the disclosures of which are incorporatedherein by reference. In this system application. the blending of dynamicbrake effort and friction brake effort takes place in the brake commandcontrol center 29, rather than in the control valve 45. Control valve 45supplies a regulated pneumatic pressure output to conduit 17 whichvaries inversely with the amount of dynamic braking effected by dynamicbraking means 28. The porportional control valve 34 supplies thepneumatic control pressure to the two end trucks of the vehicle viaconduits 17a and 17b.

The center truck braking system employs a similar proportional controlvalve 46 to provide graduated applications of pneumatic brake pressureto the pneumatic to hydraulic convertor 39. The proportional controlvalve 46 receives an input pressure from conduit 61b. and supplies anoutput pressure to a conduit 18 to the center truck. The center truckhas its own command center 31, which supplies a friction-only brakingsignal to the proportional control valve 46. This signal is derived fromthe brake command control center, and takes into consideration thevehicles total load, weight and speed, but does not take intoconsideration the amount of dynamic brake feedback received from thedynamic brake means at the front and rear trucks. Thus the proportionalcontrol valve 46 may supply substantially more pneumatic pressure to thepneumatic to hydraulic convertor 39, than does the proportional controlvalve 45 to the pneumatic to hydraulic convertors 38 and 40. Theproportioning action of the control valve includes a supply and exhaustvalve that takes the input pressure from input line 16b and eitherexhausts the pressure to atmosphere through an exhaust port, or to thebrake system for application of the friction brakes.

The proportioning control subsystem begins with the main pneumaticsupply line 16 which exits from the system supply reservoir and thefriction brake cut-out valve 63a. As illustrated in FIG. 1, the pressureis regulated by a pressure regulating valve 8 and a relay valve 9. Theregulating input line is maintained at the full supply pressure in thesupply reservoir 49. The regulating valve establishes an output pressureindicative of the desired pressure level for the system. In thepreferred embodiment this pressure level is pounds per square inch. Theoutput pressure at 8b is connected to the pilot control for relay valve9. Relay valve 9 establishes this pilot pressure on the output conduit16d. The pilot pressure at 16d is monitored by a pressure switch 15which is in turn connected to an annunciator in the brake commandcontrol center.

Referring to FIG. 2, the incoming pressure is illustrated in the upperrighthand corner of the drawing as pressure input line 16. A branchconduit 8a leads to the inlet port 7 of the pressure regulating valve 8.Pressure regulating valve 8 established a predetermined pressure, i.e.100 pounds per square inch, at the outlet port 6 and the pilot controlline 8b. This pilot pressure is communicated to the inlet port 401 ofrelay valve 9. Relay valve 9 is a conventional pneumatic relay valvewhich comprises a pilot diaphragm 402, a mainline inlet and exhaustvalve 403, and a pilot supply and exhaust valve 404. As the pilot linepressure from input port 401 pressurizes chamber 405, the pilotdiaphragm 402 and its backup piston 406 are driven upwardly asillustrated in FIG. 2 thereby displacing the main control rod 407. As407 is driven upwardly. it lifts the main mainline exhaust chamber 413and the exhaust conduit 411. Simultaneously, a portion of this pressureis routed to the pilot control assembly through interior passageway 414.The output pressure present on conduit 411 passes through passageway 414to the pilot balancing chamber 415. When the pressure in chamber 415 hasreached the pressure in chamber 405, spring means 415 returns controlrod 407 to its original position and allows the mainline control valve408 to seat on valve seat 409. This closes any further communicationbetween the mainline inlet port 410 and the mainline output port 41 1.The pilot control valve 404 is used to regulate or adjust the mainlineoutput valve 403 in response to pressure variations on the pilot controlpassageway 401.

The pressure thus communicated to the output line 16a is equivalent tothe pilot pressure established by the pressure control valve 8 on thepilot pressure line 817. If the pressure is varied, the output pressureon line 16a is also varied.

The incoming pressure to control valve 45 is therefore established inthe preferred embodiment of the invention at approximately 100 poundsper square inch. The proportional control valve used for valves 45 and46 is essentially the same, and in FIG. 2, one of the valves has beenillustrated in block form, while the other is illustrated in crosssection.

Each of the blending control valve means employed in this systemincludes four main components. Ths first component is an electricaltorque motor 134. Motor 134 responds to electrical signal variations incontrol line 19. The torque motor exerts a proportional torque oncomparator shaft 137. The second component is a pneumatic torque motor138 which applies to shaft 137 a resisting torque which decreaseslinearly with increases in the pressure applied to the friction brakesystem through output line 139a. The third component comprises a pilotvalve assembly or pressure transducer 140 which is driven by shaft 137.It serves to control the pilot pressure in a pair of pilot passages 141and 142. The fourth component comprises the supply and exhaust valveassembly 143 which serves to regulate the input pressure from manifoldline 16a to the friction brake system in accordance with the pilotpressures produced in lines 141 and 142.

The electrical torque motor 134 is of known design and comprises apermanent magnet rotor which rotates within a wound stator. Thedirection in which the motor rotates depends upon the direction ofcurrent flow through the stator, and the motor circuits are socorrelated that it always rotates in the same direction. The torqueoutput of the motor is directly proportional to the magnitude of thecurrent and the sine of the magnetic angle between adjacent unlike polesof the rotor and stator.

The output of these torque motors is connected directly to pneumatictorquc motor 138. The operation of this torque motor is fully explainedin my previous US. Pat. No. 3,536,361, the disclosure of which isincorporated herein by reference. This torque motor provides acountervailing or balancing torque on shaft 137. When the pressure inoutput line 139a is at Zero, pneumatic torque motor 138 is applying aspring loaded maximum torque output to shaft 137. Conversely. when afull service application is made, torque motor 138 applies a minimumtorque. The force exerted on shaft 137 varies inversely with pressure,and is a negative function of transducer output pressure.

. Pneumatic transducer 140 uses the combined output on shaft 137 to varythe pilot pressure supplied to the main supply and exhaust valve 143.The pneumatic transducer receives incoming pilot pressure from line 144.The pneumatic transducer responds to torque.

input on shaft 137 to provide three output conditions:

a. In the normal or application position. the transducer vents pilotpressure from the pilot passages 141 and 142.

b. At the extreme limit of angular rotation of shaft 137, pilot pressureis supplied to passages 141 and 142 and the transducer is in itsreleased position.

0. In an intermediate position between a and b above, pilot passage 141is pressurized and pilot passage 142 is vented. Any angular rotation outof this intermediate position, in the direction toward the limit oftravel, will establish pressure in pilot passage 142, and operate therelease valve to decrease pressure output. Rotation in the oppositesense, that is, toward the normal position will cause both pilot linesto be vented.

The main supply and exhaust valve 143 includes poppet type supply andexhaust valves 145 and 146 arranged to control flow from the main inletinput manifold 62 to the main brake control line 139 or to exhaust port148 and the surrounding atmosphere. Supply and exhaust valves 145 and146 are carried by spool portions a and 15017 which reciprocate inaxially aligned bores and are arranged so that if either valve (145 or146) moves in either direction, it engages the other, and causes it tomove in a valve closing direction. The opposite ends of each spool haveequal cross sectional balancing areas and interconnected by passagesextending through t he spools. This renders both valves insensitive tochanges in transducer outlet pressures. The supply and exhaust valves145 and 146 are biased closed by compression spring 149 which isinterposed between them. Each of the valves 145 and 146 may be shiftedin its respective opening direction by first and second pilot motormeans comprising compression spring 150 and diaphragm 151 (for valve145) or by compression spring 152. and diaphragm means 153 (for valve146). It should be noted that the corresponding parts of the two pilotmotor means are reversed so that in one case the valve (145) is openedby the diaphragm motor. The arrangement of the parts is such that:

a. Spring 150 opens supply valve 145 and holds exhaust valve 146 closedwhen the pilot passages 141 and 142 and diaphragm motor chambers 154 and155 are vented.

b. Diaphragm motor means 153 will open exhaust valve 146 and hold supplyvalve 145 closed when both diaphragm motors are pressurized.

c. Spring means 149 will close both the supply and exhaust valves whendiaphragm motor chamber 154 is pressurized and diaphragm motor chamber155 is vented.

These three conditions of the supply and exhaust valve 143 corresponddirectly to the three positions of 1 1 the pneumatic transducer 140referred to above.

Application of the dynamic brake and the friction brake may beaccomplished through the brake command center 29. In the preferredembodiment of the invention, control circuitry 9 carries a constantvoltage of 10 volts. A reduction in the voltage carried on control line9 will effect a brake application. The -10 volt signal on line 9 isamplified to a 020 volt signal by amplifiers 30 and 30b. The applicationis proportional to the reduction in voltage below the 10 volt standard.For example, if the control circuitry voltage was zero, a full brakeapplication would result. On the other hand, a 5 volt reduction wouldproduce a proportionally smaller brake application. In the embodimentthat is illustrated in FIGS. 1 and 2, the friction brakes for the endtrucks are controlled by proportioning valve 45 which serves to adjustthe effective braking force of the friction brakes to that of thedynamic brakes. The control sequence operates essentially in an additivemanner from the control system signal present on line 9. Thus, if thecontrol system voltage were reduced to zero, but the dynamic brakingmeans indicated a volt output from the dynamic brakes 28, the commandcenter 29 would not call for friction brake application. On the otherhand, if in the above example the dynamic brake voltage output was only8 volts, the net voltage reduction of 2 volts would produce a call for afriction brake application of approximately This would be the amountrequired to produce a fully effective blended service brake application.

It should be noted that control line 9 is interconnected to the othercars of the train to effect uniform reduction and uniform braking foreach of the cars. This may be accomplished in any one of several ways. Asingle control line may extend throughout the train, and supply theoperating control signal for each of the brake systems for each of thecars. In an eight car train, there would normally be 16 proportionalcontrol valves drawing current from the control line 9. Therefore, it isdesirable to utilize DC amplifiers 30 for each control valve which inturn supplies the operating voltage for torque motor 134.

When the brake system of FIG. 2 is in use, and propulsion controller isset in a motoring position, the circuits of the dynamic brake controller28 will be in the motoring configuration and the voltage from thecontrol center 29 will be at its maximum. The torque output of motor 134will also be at its maximum and will maintain the comparator shaft 137in a release position for transducer 140. All of the components of thesupply and exhaust valve 134 will assume their illustrated position andthe brake cylinder line 139 will be vented to atmosphere through port148.

In order to apply the brakes, the operator shifts the propulsioncontroller to a coast position, thereby reducing the voltage in controlline 9, and switching the circuits of the brake 28 to a brakingconfiguration to establish a dynamic braking effort. Since the tractionmotors now act as generators, they supply voltage to brake commandcenter 29. If the dynamic brake itself can satisfy the braking command,the signal output from the command center 29 will remain constant, andthe comparator shaft 137 will remain in the extreme release position. Inthis case, the pneumatic transducer 140 will remain in a releaseposition with pilot lines 141 and 142 pressurized and brake cylinderline 139 vented. If on the other hand, the dynamic brake is incapable ofsupplying the braking effort called for by control center 29, thereduction in the torque output of motor 134, and the pneumatic torquemotor 138 will rotate comparative shaft 137 towards the apply position.As shaft 137 rotates towards the apply position, pneumatic tranducer 140will vent the pressure present in pilot lines 141 and 142, andconsequently the fluid pressure present in chambers 154 and 155. Whenthis pressure is vented, spring means 150 will open valve 145, andsimultaneously close exhaust valve 146. When valve 145 is moved fromseat 145a it opens communication between inlet chamber 156, and outletchamber 147, establishing communication between input manifold line 62and brake line 139. Since the exhaust valve 146 has been closed, arpressure will now be supplied to the friction braking means throughbrake line 139. As the pressure in 139 and 139a develops, the torqueoutput of pneumatic torque motor 138 will be reduced, and the torquewhich it applies to comparator shaft 137 will be reduced. Accordingly,as the braking effort of the friction brake means approaches the levelrequired to compensate for the deficiency in the output of the dynamicbrake means, the electrical torque motors 134 will begin to rotatecomparator shaft 137 towards the intermediate or lap position. When thesum of the outputs of the friction and dynamic braking means is equal tothe selected braking effort, the torques exerted on shaft 137 will bebalanced, and pressure transducer 140 will rest in an intermediate orlap position.

In the lap position, pilot passage 141 is pressured while pilot passage142 is vented. As the pneumatic transducer moves to its lap position,the working pressure in chamber 154 will increase, and the diaphragmmotor means 151 will overpower spring allowing spring means 149 to closethe supply valve 145.

After train speed has been reduced to a low level, the braking effort onthe dynamic brake means will begin to fade". This will reduce the amountof voltage supplied to the command center 29. When the train enters thisportion of the braking cycle, the torque acting on comparator shaft 137will again become unbalanced in the opposite direction and pneumatictorque motor 138 will shift the pressure transducer to an applicationposition. This position will vent both of the working spaces 154 and 155through pilot control lines 141 and 142 and allow spring means 150 toopen supply valve 145 and close exhaust valve 146. Air under pressurewill now be supplied through inlet chamber 156 and exhaust port 147 tobrake line 139. As the pressure rises in brake line 139, the increasedpressure will effect a reduction in the torque output of pneumatictorque motor 138. When the braking effort of the pneumatically operatedfriction brake is increased sufficiently to offset the decrease in theoutput of the dynamic brake due to fade the torque motors 134 and 138will return the pneumatic transducer to a lap position.

In view of the foregoing discussion, it should be evident that,regardless of the effect of speed on dynamic braking effort, the systemwill always graduate the friction braking effort as needed to maintainthe total braking effort required.

As was pointed out previously, the center truck is also braked by meansof pneumatically controlled friction brakes. The proportional controlvalve 46 is identical to the proportional control valve 45, and operatesin essentially the same manner as heretofore described for control valve45. However, the brake command control center provides a decreasingcontrol signal on electrical line 20 that is independent of the dynamicbrake output. The control signal generated by the center truck brakecommand 31 will vary according to the speed of the train, and thevehicle load weight, but will not vary in accordance with dynamic brakeoutput. When the voltage reduction is impressed on control line 9, itwill be transmitted without mixing to the center truck command 31, andthereafter to amplifier 30b and control valve 46. During the dynamicbraking effort, the two end trucks may be applying only to 30% of thetotal air pressure available to the friction brake, while the centertruck is applying a 100% braking effort to the friction brakes.

The Variable Load System The variable load system of the presentinvention utilizes a separate air spring for each truck, and a separatevariable load valve interposed between the proportional control valves45 and 46, and the pneumatic to hydraulic convertors 38-40. As such, thevariable load system is an independent intervening system whichregulates the maximum amount of fluid pressure which may be supplied tothe convertors 38-40.

FIG. 3 illustrates in schematic and cross sectional form a portion ofthe variable load system of the present invention. The components of thevariable load system are fully described in US. Pat. No. 3,730,597entitled Variable Load Rate Control Apparatus assigned to the assigneeof the present invention, the disclosure of which is incorporated hereinby reference. The variable load system is regulated by the air springs51a, 51b, and 510 which are illustrated schematically in FIG. 3 by airspring 51. The air spring 51 has been found to have superior ridecharacteristics to the coil spring, and supports the vehicle body on itsrespective truck. As the load of the vehicle increases, the air pressuremaintained in air spring 51 is increased. This increase in air pressureis provided by piping 72 and a regulator. As the load of the vehiclegoes up, the air pressure in the air spring must also go up to maintainthe car body at a constant height. The variance in air pressure in airspring 51 between an empty load and full load is used to regulate thevariable control valve 53.

The air spring mechanism 51 is usually in the form of one or moreflexible bags located between the car body and the trucks. Theregulating valve mechanism (not shown) is provided which is actuated inresponse to vertical movements of the car body caused by variations inthe load of the car body to selectively increase or decrease thepressure in air spring 51. Thus, if the load on the car body increases,the car body moves downwardly towards the trucks and actuates theregulating valve mechanism (not shown) to increase the pressure in airspring 51. This increase in pressure then raises the car body back toits original predetermined height above the platform or rails.Conversely, if the load on the car decreases, the car body rises withrespect to the trucks and actuates the regulating valve mechanism todecrease the pressure in the air spring 51 which in turn lowers the carback to its original predetermined height.

The variations in pressure present in air spring 51 are used to regulatethe variable load control valve 53 illustrated in FIG. 3. As illustratedin FIG. 1, in coming air from the main reservoir passes through checkvalve 50, and enters supply reservoir 49 and air spring 51 throughconduit 72. The output of the proportional control valves 45 and 46 issupplied via manifold 17 and 18 to the variable load valves 53a, 53b,and 53c,

14 one of which is illustrated in FIG. 3. The incoming air pressurepresent on manifold line 72 is piped in a parallel manner to the threevariable load valves.

The variable load valve receives a pilot pressure from the air springvia conduit 75, which enters valve 53 through inlet passage 76.Simultaneously, operating pressure from the proportioning valve 45 issupplied to the relay portion of variable load valve 53 through entryport 77, and is conveyed to the magnetic dump valve 59 via outputmanifold line 78, and conduit 79. Magnetic dump valve 59 is mounted tothe housing of convertor 38, for reasons to be discussed; thus conduit79 is also shown in phantom to convertor 38 in FIG. 4. In the preferredembodiment of the invention, the variable load valve is mounted on theexterior of the pneumatic to hydraulic convertor 38. An exploded viewformat has been used in FIGS. 3, 4 and 5 to clearly illustrate theinterrelationship of the various components.

Variable load valve 53 is illustrated in an empty configuration. Thatis, the minimum amount of pressure is presented to the pilot portion ofvalve 53 via air spring 51. This minimum air pressure is conveyed to thevariable load valve through conduit 75, inlet port 75, an inletpassageway 81, to the pilot chamber 82. This pilot pressure togetherwith spring means 83 exerts a force acting in the direction of arrow Aon the top of piston 84. Balancing this force is a force acting upwardlyon the bottom of piston 84 exerted by spring 104 through valve body9293, retainer plate 86 and.

spring means 85. This same force is also exerted by spring on retainerplate 86. Since the opposing forces of spring 85 are balanced, thepiston is at rest at a predetermined location within the guide bore 87.

When the car is empty, the pilot pressure in chamber 82 is at itsminimum, and the piston 84 will rest with its shoulder portion 88 inimmediate abutment with the shoulder 89a of upper cap member 89. If thepressure in chamber 82 were at its maximum as in a fully loadedcondition, the piston 84 would be driven in the direction of arrow A tocause the retainer plate 86 and control diaphragm 90 to abut theshoulder 91a of lower cap 91.

The relay portion of the variable load valve comprises supply valve 92,exhaust valve 93 and their associated valve seats 94 and 95. Asillustrated in FIG. 3, supply valve 92 is open and air pressure enteringinlet passageway 77 passes into chamber 97. From chamber 97 it entersthe relay valve through transverse port 98 and passes between the supplyvalve 92 and its seat 94 to coaxial bore 99. From bore 99, it entersoutput chamber 100 and exits through output port 78. Air pressure inexhaust passage 100 will simultaneously traverse through interior port101'to the control chamber 102 immediately below diaphragm 90.

The inlet and exhaust valves 92 and 93 are linked together by means ofpin 103. As illustrated in FIG. 3, exhaust valve 93 is firmly seatedagainst exhaust seat by means of spring 104. At the same time, theinterconnecting pin 103 has lifted input valve 92 from valve seat 94 toopen the relay portion of the valve.

As the air pressure from the proportioning valve 45 passes through therelay valve portion of variable load valve 53, it pressurizes chamber102 until the pressure therein is sufficient to compress springs 83 and85 and move diaphragm 90 and retaining means 86 upwardly as indicated byarrow B. As retaining means 86 is moved upwardly, exhaust valve 93 isdriven upwardly by pin 103, supply valve 92 and spring 104. When thepressure in chamber 102 has reached a predetermined level, supply valve92 will seat at 94 and close coaxial bore passage 99.

As the vehicle is loaded, the air pressure present in air spring 51 willincrease, and consequently the air pressure present in conduit 75, inletport 76, inlet passage 81, and pilot chamber 82 will also increase. Asthis pressure increases, the pressure on piston 84 and the force ofspring 83 will combine to move piston 84 downwardly in the direction ofarrow A. As piston 84 moves in the direction of arrow A, it will movespring means 85 and retaining means 86 downwardly. Retaining means 86will then unseat supply valve 92, allowing additional pressure to betransmitted to relay valve 52 and control chamber 102. Eventually,piston 84 is moved downardly, the retaining means 86 will move intoabutment with shoulder 91a of end cap 91. Any subsequent increases inpressure will only result in a further compression of spring 85. Itshould be noted that at this point, the pressure required in chamber 102to move the diaphragm 90 and retaining means 86 up wardly as indicatedby the arrow B will be substantially greater than it was for the emptycar. The pressure in chamber 102 must rise to a force sufficient toovercome combined forces of spring 85 and/or spring 83 and the pressurein chamber 82 to reseat the supply valve 92. When such equilibrium isreached, the supply valve 92 will again close, shutting off passage 99.

It should be noted that the spring constants of springs 83 and 85 arevery important. As discussed previously the pressure in air spring 51varies according to a first relationship determined by the ratio of theload to the weight of the car body only, and not the rail vehicleitself. The pressure delivered to the convertor 38 varies however,according to a second and different relationship determined by the ratioof the load to the total weight of the vehicle. Accordingly, fullbraking pressure need not vary as greatly as the air spring pressure. Inthe preferred embodiment of the invention, the effective area of piston84 and diaphragm 90 are substantially equal. The spring constant spring85 however is larger than that of spring 83 by an amount which causesthe full service brake pressure to vary in accordance with the secondrelationship, even though the air spring pressure varies according tothe first relationship.

It should be pointed out that the pressure in air spring 51 willincrease the decrease incrementally at each stop as passengers enter andleave the vehicle.

If the load of the vehicle is decreased, regulating means (not shown)will exhaust air from the air spring means 51. When pressure is reducedin the air spring 51, it is correspondingly reduced in pipe 75, inletpassage 76, and pilot chamber 82. As the pressure is reduced, the pistonmeans 84 will move upwardly as indicated by the arrow B and the controlpressure already present in control chamber 102 will move the diaphragm90 and retaining means 86 upwardly as indicated by the arrow B. Sinceexhaust valve 93 is retrained in a fixed position by means of linkingpin 103, and the seating of valve 92 upon valve seat 94, exhaust valve93 will be upseated from its valve seat 95, and the excess pressure incontrol chamber 102 will be allowed to flow through the axial bore 105into chamber 106 and out exhaust passage 107. This also serves to reducethe pressure in passageway 101, exhaust passage 78, manifold lines 79and pneumatic to hydraulic convertor 38. When the control chamber 102has been sufficiently vented, the spring 85 will overcome the pressurein chamber 102 and will again move retainer means 86 downwardly asindicated by the arrow A, and cause valve seat 95 to engage exhaustvalve 93. As can be seen from the foregoing description, variable loadvalve 53 produces a variable pressure that is partially proportional tothe pressure present in air spring 51, and the load carried by thevehicle. This difference in spring constants between spring 83 andspring is provided to provide for the difference in load to weightratios between the load and the car body as opposed to the ratio betweenthe load and the weight of the vehicle. These pressures areproportionally reproduced in the brake cylinder by means of thepneumatic to hydraulic convertors.

The foregoing description has centered around the variable load systemfor the two end trucks of the vehicle. The same interrelationship existsfor the center truck and control valve 46. The proportioning controlvalve 46 provides a variable output pressure which may in turn bemodified by the variable load valve 531). The operation of this valve isidentical to the operation of valve 530 hereinbefore described indetail.

The Pneumatic to Hydraulic Convertors FIG. 4 is a cross-sectioned anddiagrammatical representation of the pneumatic to hydraulic convertors38-40. This subsystem comprises a pneumatic to hydraulic convertorgenerally designated as 38, the pneumatic control input line 79, thedump valve 59, a hydraulic motor 220 and a hydraulic slack adjuster 41.Converter 38 is disclosed in copending application Ser. No. 417,707,filed Nov. 20, 1973, now abandoned but continued-in-part in applicationSer. No. 501,939 filed Aug. 29, I974.

The pneumatic to hydraulic convertor 38 includes a fluid responsivepneumatic motor means. This motor means comprises a working chamber 214which is defined by the booster housing 215 and the flexible diaphragmmember 216. The flexible diaphragm 216 is backed by a reciprocatingpiston 217 which is fixably attached to a reciprocating connecting rod218. As illustrated in FIG. 4, when working chamber 214 is pressurized,the flexible membrane 216 and the working piston 217 are displaced tothe left thereby actuating a hydraulic motor means generally indicatedby the numeral 220. The reciprocating working piston 217 is biased tothe position illustrated in FIG. 4 by virtue of a resilient spring means219 located within the housing of the pneumatic to hydraulic convertor.Housing member 215 also defines a hydraulic sump 215a for storage ofhydaulic fluid for the hydraulic motor means 220. Housing member 215also defines an integral dump chamber 215b and a pair of dumppassageways 221 and 222. Since only a portion of the passageways 221 and222 is illustrated in the cross-sectional portion of FIG. 4, theremaining portion of the passageway is indicated by the dotted lines221a and 222a. See the FIG. 6.

The interconnection between the working chamber 214 of the fluidresponsive pneumatic motor means and the dump chamber 215b is normallyclosed by means of magnetic dump valve 59 which is interposed betweenpassageways 221 and 222. This interconnection and the operation of thedump valve 59 will be hereinafter later described.

The pneumatic to hydraulic convertor illustrated in FIG. 4 also includesa slack adjusting means generally designated by numeral 41. Inoperation, air under pressure is supplied through conduit 79 to thefluid responsive pneumatic motor means 214. As the pressure in workingchamber 214 increases, it will drive diaphram 216 and piston member 217to the left as indicated by arrow A in FIG. 4. The force exerted by theflexible cliaphragm and piston is transmitted through the connecting rod218 to hydraulic piston 226 in the hydraulic motor means 220. Thehydraulic piston 226 in turn supplies hydraulic fluid under pressure tothe slack adjustor 41. The output of slack adjustor 41 is transmittedthrough hydraulic line 160 to the hydraulic actuators mounted on thedisc brakes.

When the control valve 45 has received a signal to de-energize thefriction brake subsystem, it will vent control line 17 to reduce thepressure to the variable load valve and conduit 79. As the pressure inconduit 79 and chamber 214 is reduced, the spring means 219 will returnpiston member 217 to the extreme righthand position illustrated in FIG.4. As the piston member is withdrawn, the connecting rod 218 returns thehydraulic piston 226 to its extreme righthand position as illustrated inFIG. 4.

The Wheel Slip Dump Valve and Chamber As was previously indicated, thepresent invention includes an independent intervening subsystem fordetecting and correcting wheel slip. This subsystem is designed todetect differences in speed between axles, and to detect angulardeceleration during synchronous slips. The system provides for a rapidreduction in the pneumatic brake cylinder pressure of the sliding truckduring brake application when a slip or slide is detected. The system isengineered to provide for failsafe operation through the use of normallyde-energized relays and normally deenergized magnet drives which operate the dump valves. The system is made up of three basic subsystems;a detection system, relaty logic, and the magnetic dump valve assemblyillustrated in FIGS.

The dump valves 59-61 used in the present invention are high-capacity,fast-response, normally closed magnet valves. They are installed in thehousing of the pneumatic motor means and normally close a passagewayextending between the pneumatic motor means and an integral dumpchamber. As illustrated in FIG. 4,

the dump chamber 2115b is defined by the walls of casing 215 as anintegral part thereof.

The present invention uses large high-capacity passageways 221 and 222to interconnect the working chamber 214 with the dump chamber 215a.These passageways, together with the high capacity dump valve 59,provide almost instantaneous communication between working chamber 214and the dump chamber 215b. This vastly reduces the response time of thepresent invention. See FIG. 6.

The dump valve, dump chamber, and the associated passageways between thedump chamber and the pneumatic motor means are illustrated in FIGS. 4-,5 and 6. FIG. 4 illustrates the dump chamber 215b in cross sec tion,with the dump valve 59 arranged immediately below the chamber to provideclose communication between the dump chamber 215]; and the workingchamber 214. The dump chamber 21519 is normally exhausted through anexhaust orifice 247 which provides a restricted through passageway fromchamber 2l5b to atmosphere.

FIG. 6 illustrates the dump chamber 215!) and the interconnectingpassageways 221 and 222 in isometric section. As illustrated in FIGS. 5and 6, passageway 1 221 extends from the dump chamber 215]) to a centralpassageway 260 which houses dump valve 59 and interconnects the slottedpassageway 221 with slotted pas sageway 222 when valve 59 is actuated.Casing member 215 also defines a series of threads 261 in the innerperiphery of passageway 260 to engage the dump valve 59. This engagementis more fully illustrated in FIG. 5. FIG. 5 illustrates the dump valveand passageways 221 and 222 in cross section. The dump valve 59 is ahighcapacity, fast-response two-way normally closed magnet valve. It isrotatably threaded, as illustrated in FIG. 5, into threads 261 to closecommunication between passageways 221 and 222. Dump valve 59 includes anaxially reciprocating spool means 262 having a pair of valve discs 263and 264 mounted at the end thereof. These valve discs 263 and 264reciprocate between a first, valve closing seat 265 and a second, valveopening seat 266. Valve disc 263 is normally urged into contact withvalve seat 265 by means of resilient spring 267. Dump valve 59 is thesubject of application Ser. No. 484,188, filed June 28, 1974.

The dump valve 59 is opened by application of an electrical controlsignal to the electromagnetic winding 270 which surrounds areciprocating armature 27. When the electromagnetic coil 270 isenergized, it displaces the movable armature 271 to the right asillustrated in FIG. 5, thereby displacing spool 262 to the right andcompressing spring 267. As spool 262 is moved to the right, valve disc263 is unseated from seat 265 thereby opening communication between theworking chamber 214, passageway 222, valve inlet passageway 272, theannular seating chamber 273, the axial passageway 275, and the exhaustpassageways 276. Exhaust passageways 276 exit into the dump valvepassageway 221 defined in casing 215 and to dump chamber, 2151:. At thesame time disc 264 is moved to the right against annular seat. 266 thussealing off air supply from control valve 45 on line 79. The air inchamber 214 initially expands rapidly into chamber 215b, effecting arapid reduction in local brake pres sure in response to a wheel slipsignal and then escapes more slowly via orifice 247, preventing completepres sure loss in chamber 214. When the electromagnet 270 isde-energized, the return spring 267 will unseat valve disc 254 from seat266 and engage valve disc 263 with valve seat 265 to thereby closecommunication between annular seating chamber 273 and exhaustpassageways 276. This admits air from line 79 and seals the workingchamber 214 of the pneumatic motor means thereby preventing anyadditional pressure drop therein. Thus small adjustments in the pressurein chamber 214 may be made to account for wheel slippage due to changingtrack conditions; whereas, complete depressurization of chamber 214 isnot required. This system is discusssed more fully in my co-pendingapplication Ser. No. 417,707 for Pneumatic to Hydraulic Convertor withIntegral Dump Chamber, filed Nov. 20, 1973, which is incorporated hereinby reference.

The hydraulic motor means of the present invention includes thehydraulic piston member 226, a hydraulic cylinder 227, and a hydraulicworking space 228. Hydraulic fluid is stored in the reservoir 215a whichis defined by the external walls of casing member 215. As illustrated inFIG. 4, the hydraulic cylinder 227 is a separate cylindrical memberwhich is inserted into the casing member 215 and secured thereto by thehydraulic slack ad justor and mounting bolts 23,0. Hydraulic workingfluid enters the hydraulic working space 228 through a first workingport 231 to insure that the working space 228 is completely filled withhydraulic fluid at all times. A second fluid port 232 is provided whichcommunicates with a second working space 233 that cooperates with piston226 to transfer additional hydraulic fluid to the working chamber 228when a slack adjustment has been made by slack adjustor 41. Thistransfer of fluid during slack adjustor will hereinafter laterexplained.

Piston member 226 also includes a check valve generally indicated by thenumeral 235. The check valve 235 is used to close a passageway 236 whichextends through piston member 226. Check valve 235 normally closescommunication between the working chamber 228 and the second chamber233. Passageway 236 allows the transfer of hydraulic fluid from workingspace 233 to wroking space 228 when a slack adjustment is made by theslack adjustor 41.

The slack adjustor 41 comprises a stepped cylinder 237 and adifferential area reciprocal piston means 238. This piston means isformed with a first alrge diameter piston portion 238 and a secondsmaller diameter piston portion 240 which fit into the correspondingportions of stepped cyliner 237. The piston means 238 divides thecylinder 237 into a pair of working spaces 241 and 242 with the firstworking space 241 being at the smaller diameter end of the cylinder, andin continuous communication with the working chamber 228 of thehydraulic booster 38. The two working spaces 241 and 242. are normallyisolated from each other, but under certain conditions, fluid can betransferred between the spaces through a valve unit 243 installed withinthe differential piston 238. Under normal working conditions, the valveunit 243 is closed by the hydraulic pressure present in chamber 241, andby its own internal compression spring 244. The valve unit may, however,be unseated by excess pressure present in working space 242. The valvethereby performs a check valve or release function and allows flow fromspace 242 to space 241 when the pressure in the former exceeds that ofthe latter by a predetermined amount. The valve means 243 can also beunseated mechanically by means of a push rod 245 which is affixed to theend of the cylinder 237. The push rod is effective to open the valvejust before piston 238 reaches the limits of its leftward travel, orapproximately l/16th of an inch before the piston abuts the endcap. Thestroke of piston member 238, and the respective volume of chamber 242determine the clearance between the brake pads and discs.

When the pneumatic to hydraulic convettor is in service, and the brakesare released, the hydraulic motor 220 and the slack adjustor 41 assumetheir illustrated positions in FIG. 4. When a service brake applicationis initiated, the pneumatic piston 217 is displaced to the left ashereinbefore previously described. As the piston 217, connecting rod218, and hydraulic piston 226 are displaced to the left, port 231 isclosed and working chamber 228 is pressurized. As chamber 228 ispressurized, the hydraulic fluid in chamber 241 is also pressurized, andthis pressurization displaces piston member 238 to the left displacinghydraulic fluid from the work ing space 242 to the hydraulic brakeactuators via conduit 160. If the path clearance is less than that whichthe slack adjustor 41 is designed to maintain, the brake pads (FIG. 7)will be moved into contact with their respective brake discs beforepiston member 238 reaches the limit of its leftward movement. At thispoint, the

20 pressure in working space 242 will rise above the pressure in space241 as a result of the difference between the cross sectional areas 'ofpiston portions 239 and 240. When the pressure differential reaches thesetting established by valve unit 243, the valve will be unseated topermit flow of hydraulic fluid from chamber 242 into chamber 241. As aresult, the piston member 238 will be shifted all the way to itslimiting lefthand position in immediate abutment with the endcap ofcylinder 237.

Just before the piston member 238 abuts the endcover, a push rod 245will open valve means 243 and the hydraulic motor 226 will be in directcommunication with the hydraulic actuators of the disc brakes. Thepressure developed in the various working cham bers 228, 241 and 242,and in the hydraulic actuators will be proportional to the pneumaticcommand pressure initiated by the control means and presented to thepneumatic motor means and working space 214.

When the service brake is released, the output pressure developed by thecontrol means is exhausted and the spring means 219 returns the piston217, thereby withdrawing the hydraulic piston 226 to reduce thehydraulic pressure in working spaces 228, 241 and 242. Accordingly, thepressure now present in working chamber 242 will be effective to shiftpiston 238 to the right to effect closure of valve means 243. As thehydraulic means 226 retracts and withdraws hydraulic fluid from theworking space 228, the slack adjustor piston 238 will also shift backtowards the initial position and in effect transfer hydraulic fluid fromworking space 241 to working space 228.

If however, increased wear on the friction brake pads has resulted in ashoe clearance which is initially greater than desired, the slackadjustor functions in the following manner. The slack adjustor piston238 will reach its lefthand position abutting the end cap of cylinder237 before the hydraulic actuators have brought the brake pads intocontact with the disc. At this point, the push rod 245 will unseat thevalve means 243 so that the additional hydraulic fluid required to takeup the remaining shoe clearance, and to thereby develop the desiredlevel of braking force can be transferred through valve means 243 to thehydraulic actuators. Since a brake application has caused the pistonmeans 238 to move full stroke into engagement with the end cap ofcylinder 237, it follows that the subsequent release of the servicebrake will cause piston 238 to withdraw from the hydraulic actuatorsexactly the required quantity of hydraulic fluid. As stated previously,the volume of chamber 242 and the stroke of piston 238 is designed toeffect the withdrawal of fluid establishing the proper clearance betweenthe brake pads and the brake disc. Thus it follows that if the shoeclearance is additionally too great, slack adjustor 41 will reduce it tothe desired value.

In each of the foregoing situations, the amount of hydraulic fluidreturned to the working space 28 and reservoir 215a must be altered.When the shoe clearance is initially too great, the quantity of oildischarged from chamber 228 during the application will necessarily begreater than the quantity returned when the brakes are subsequentlyreleased. Therefore, during the release, the slack adjustor piston 238will return to its initial position and contact abutment 246 before thehydraulic piston 226 has reached its retracted position. In thissituation. oil will be transferred from the working chamber 233 toworking chamber 228 through inclined passageway 236 and check valve 235.In situations where 211 it is desired to use the booster suctioninitiated by the withdrawal of piston 226 to augment the retractionforces acting on the hydraulic brakes, the degree of assistance can beincreased or decreased by reducing or increasing the diameter ofpassageway 236 and the design parameters of check valve 235.

The Hydraulic Brake Actuators FIG. 7 is a partially cross sectioned viewof the emergency and handbrake control system together with disc brakeacutator 32. This apparatus is disclosed in copending application Ser.No. 480,439 filed June 17, 1974', and Ser. No. 480,440, filed June 17,1974. In the preferred embodiment of the invention, the brake systememploys disc brakes and hydraulic acutators. Although it would bepossible to use pneumatic actuators and a conventional brake system, theuse of hydraulic actuators and disc brakes provides precise control thatis needed for a fast response brake system. It also provides asubstantial reduction in the space required for mounting the cylinders.As illustrated in FIG. 7, each of the wheels is equipped with asegmented disc illustrated by disc 314 and 315. The brake members aremounted on the vehicle truck and define a pair of link arms 382 and foreach brake member. Each of the link arms carry friction pads 324 and 325which bear against the disc 314 and 315.

The friction brake illustrated in FIG. 7 is actuated by a firsthydraulic motor 321. The working space of motor 321 exerts hydraulicpressure against piston member 322 and an opposing countervailing forceagainst the rear wall 351 and housing 320. These opposing forces aretransmitted to link arms 382 and 383 through pivot points 353 and 354.The opposing forces generated on either side of working space 321 arethus transmitted through the pivot points 353 anad 354 into effectivebraking force on pads 324 and 325.

The brake actuator illustrated in FIG. 7, also includes a secondhydraulic motor 327 and a spring operated motor 325. Under normaloperating conditions, spring motor 325 is restrained by hydraulic fluidin working space Upon a reduction in hydraulic pressure, the bellevillesprings 325 are allowed to exert their working pressure against piston326 and thereby actuate the. disc brake.

Although belleville springs are capable of exerting great force, theyexert it only through a very short working space. Thus it is necessaryto insure that the brake pads 324 and 325 are always in close contactwith the disc 314 and 315. It is desirable in the design of such a braketo have a working space on the order of I87 thousandths of an inch. Itis therefore necessary to provide a mechanical slack adjustor to insurethat the pressure exerted by the belleville springs is transmitteddirectly to the brake pads rather than being lost in the slack or playpresent through normal wear in the mechanical linkage.

To provide this mechanical slack adjustment, a freely rotating leadscrew 359 with a suitably steep pitch to its threads is journaled forrotation and reciprocation in bearing means 360. The lead screw 259 isthreaded into member 361 which is fixably and rigidly secured to pistonmember 322. The lead screw 359 is capable of approximately 3/l6th of aninch of axial travel and is limited in its axial travel by snap ring 352and by flange member 363. A spring means 364 is also provided to insurethat under normal operating circumstances the lead screw 359 is biasedto its leftward position wherein 22 the snap ring 362 engages bearingmeans 360. Lead screw 359 is also equipped with a conical pressure lead365. Conical head 365 engages a similar and mating surface 366 formed onthe inner periphery of piston 326.

In normal operation, working space 327 is pressurized and piston member326 is urged to a leftward position maintaining a constant bias onbelleville springs 325. When the hydraulic motor 321 is actuated, pistonmeans 322 is displaced to the right as illustrated in FIG. 7, and thisdisplacement will rotate lead screw 359 by means of member 361. Springmeans 364 is sufficiently resilient to maintain lead screw 359 within afew thousandths of an inch of its leftward position with the snap ring362 adjacent bearing 360 even while it is rotating by virtue of theforces exerted on it by member 361 and piston 322. Thus when a servicebrake application has been made, and disc brake pads 324 and 325 havebeen brought into contact with disc 314 and 315, the conical head 365will be held within a few thousandths of an inch of this relativeposition, When the pressure in working space 321 is vented, the servicebrakes are deenergized. The disc pads are free to retract to whateverdegree the separate hydraulic slack adjustors will permit.

The spacing between 365 and. 366 is also maintained under normaloperating conditions within a few thousandths of an inch. When thehydraulic fluid in motor 327 is vented, the belleville spring motor 325will displace piston member 326 to the right, urging it into contactwith the conical head 365. Once the working surfaces of 365 and 366 haveengaged one another, the lead screw 359 will be secured against anyfurther rotation. At this point, the entire force of the bellevillespring motor 325 is exerted through piston member 326, lead screw 359,member 361, and piston 322 to the link arms 382 and 383. The reactiveforces are then translated around pivot points 353 and 354 to the brakepads 324 and 325.

The slack adjustor previously described with respect to lead screw 259is a mechanical slack adjustor intended to compensate for piston motiondue to pad wear. It plays no part in service brake operation and doesnot eliminate the need for hydraulic slack adjustment, which isdiscussed regarding FIG. 4.

In the preferred embodiment of the invention, the belleville springmotor has a preferred working distance of approximately 3/16 of an inch.This working space must be carefully selected within the bellevillespring design parameters since overtravel in the compression directioncan destroy the spring, while overtravel in the extension direction willrender the spring motor ineffective. In designing this motor, it wasassumed that the fully released position for the emergency and handbrakewould be achieved with the 542 pounds per square inch of hydraulicworking pressure. This pressure is below the normal minimum servicepressure for a lightly loaded vehicle. The normal service brakingapplication for an average load vehicle was assumed to be 679 pounds persquare inch, and this pressurization on the belleville spring motor 325resulted in compression of the belleville springs totalling 0.610 inchfrom their free position, and resulted in compression of the springpiston against its stop at a load of 1 1,500 pounds per square inch.When the hydraulic motor 327 was vented, and spring motor 325 allowed toapply its fully effective braking force, the maximum overtravel in theextension direction as limited by the stop was approximately 0. l80 inchand corresponded to a load of 8,4l pounds per square inch. It is alsonecessary to provide differentially sized working areas for the firsthydraulic motor 321 and the second hydraulic motor 327. This is toinsure that the second hydraulic motor 327 would be capable of fullycompressing the springs of spring motor 325 even under lightly loadedconditions. The necessity for this will be hereinafter explained withrespect to the emergency and parking brake control system.

Once the emergency brake or parking brake has been applied, it isnecessary to apply the service brake to deenergize it. The serviceapplication fills working space 327 as will be hereinafter laterexplained, and pressurizes the second hydraulic motor to displace piston346 to the left as illustrated in FIG. 7 against spring motor 325. FIG.7 also illustrates in partial cross section a mechanical retractiondevice for the spring motor 325. As illustrated in FlG. 7, the emergencypiston 326 is equipped with internal screw threads 310. These screwthreads are engaged by a retraction screw 311 which is equipped withexternal threads 312. The retraction screw 310 is journaled for rotationin end cap 313 and is restrained from inward axial travel by means ofthrust washer 300. Retraction screw 311 is rotated by means of a socketdrive wrench which is inserted into a standard socket 305. Socket 305 isfixably secured to retraction screw 311 by means of a pin 366. Toretract the emergency piston 326 and compress spring motor 325, a wrenchis inserted into socket 305 and rotated in a clockwise manner. Threads310 and 312 then retract the emergency piston 326 to the left asillustrated in FIG. 7 thereby compressing the spring motor 325.

The retraction device also includes a second spring motor 367 which ispinned to both the retraction screw 311 as illustrated at 308 and to therear wall 313 as indicated at 309. As the retraction screw and socket isrotated in a clockwise manner, spring member 367 is wound, therebyexerting a counterclockwise torgue on retraction screw 311. However, theforce of the belleville spring motor 325 is so great that substantialfriction is generated between the threads 310 and 312 and the thrustwasher. Even though spring means 367 is exerting an unwinding bias onretraction screw 311, retraction screw 311 is prevented from unwindingby virtue of the friction exerted on screw threads 312 by the springmotor 325.

Lead screw 311 and emergency piston 326 are also equipped with abutments372 and 371 which provide a definitive stop for any further retractionof the emergency piston 326 by lead screw 311. This prevents an overcompression of the belleville spring motor 325 and subsequent jamming ofthreads 310 and 312.

The retraction device is disengaged by pressurizing the second hydraulicmotor 327. When a service application is made, or when the servicebrakes are cycled, the second hydraulic motor 327 is pressureized aswill be hereinafter described. When the pressure in the second hydraulicmotor 327 equalizes the bias exerted by spring motor 325, the frictionbetween threads 310 and 312 no longer exists, and the spring means 367is then free to rotate the retraction screw 311 in a counterclockwisemanner to its original position. If desired, the manual retractiondevice can also be released manually rotating socket 305 with the wrenchin a counterclockwise manner to restore lead screw 311 to its originalposition.

The Emergency And Parking Brake Control System The braking system of thepresent invention also includes a separate emergency and parking brakecontrol system which is capable of operating independently andredundantly on each truck of the vehicle. The emergency brake controlsystem uses a hydraulicallyrestrained spring motor which applies thebraking pads to the discs when the system is energized. These brakes maybe applied by merely venting or releasing the hydraulic motor whichrestrains the spring motor. This provides a completely failsafe mode ofoperation wherein the only common component between the emergency systemand the service brake system is the brake head and disc pads. Thissystem is also disclosed in application Ser. Nos. 480,439 and 480,440,previously discussed.

The handbrake or parking brake system also uses thehydraulically-restrained spring motor used by the emergency brakesystem. For both of the systems, a completely separate subsystem isprovided for maintaining the hydraulic pressure in the motor whichrestrains the spring motor.

The control system illustrated in FIG. 1 is equipped with a masterpressure switch 15 which indicates a failure of the pneumatic controlsystem to the brake command control center 29. In addition, each of thetrucks is equipped with a pressure switch 501, 502 and 503 which arealso connected to annunciators in the brake command control center 29.Upon noting a malfunction, the operator may elect to continue theregular service brake application, or to energize the emergency brakesystem. Additionally, the pressure switches and annunciators will alertthe operator to a brake or brake control system that has failed torelease.

As discussed previously, the control system illustrated in FIG. 1 isequipped with pneumatic to hydraulic convertors 38, 39 and 40 to supplyhydraulic fluid under pressure to the friction brake actuators 3237. Theoutput of the convertors is piped to a plurality of manifold blocks504-506 which divert the hydraulic fluid to their respective actuators.Each of the manifold blocks 504506 is equipped with a companion block504a-506a for supplying hydraulic fluid under pressure to actuators 33,35 and 37. Each of the manifold blocks 504-506 is equipped with a magnetvalve 301-303 respectively which are operated by electrical signalsimpressed upon signal control line 304. The signal control line leads tothe operators console and the operating controls for the emergency brakecontrol system 14 and the parking brake control system 10.

Each of the actuators 32-37 has a first hydraulic motor for servicebrake applications, a spring motor for emergency and handbrakeapplications, and a second hydraulic motor for restraining the springbrake. Actuator 32 is connected to manifold 504 via a service brake line511 and an emergency and handbrake control line 512. Conduit 51 1defines a first supply line for the service brake actuator, whileconduit 512 provides a second supply line for the second hydraulicmotor. Likewise, first 513 and second 514 supply lines lead to thesecond manifold 504a which supplies hydraulic pressure for actuator 33.This hydraulic pressure is supplied along service conduit 516 andemergency and handbrake conduit 517. The use of the first and secondsupply lines and their associates manifolds is duplicated for each ofthe remaining actuators 3437.

The interior piping of manifold 504 is more fully illustrated in crosssectional form in FIG. 7. The incoming supply conduit 160 supplieshydraulic fluid under pressure to junction 518. A portion of this fluidis diverted through the first supply lines 511 and 513 to the servicebrake portions of actuators 32 and 33. An additional portion of thefluid may be diverted into the emergency and handbrake control systemthrough pressure-responsive check valve 530.

As was previously described with respect to FIG. 7, actuator 32 includesa housing means 320, and a first fluid motor 321 energizing a servicebrake application. As motor 321 is pressurized, piston 324 is displacedto the right as illustrated in FIG. 7 to bring disc pads 324 and 325into engagement with disc means 314 and 315.

Actuator 32 also includes a spring motor 125 which exerts pressureagainst piston 326, conical head 365, and piston 322 to bring the discpads 324 and 325 into engagement with their respective discs 314 and315. The actuator also includes a second hydraulic motor means 327 whichurges piston 326 to the left as illustrated in FIG. 7, restrainingspring motor 325 and rendering it inoperative during service brakeapplications. Under normal operating circumstances, if the secondhydraulic motor 327 is de-pressurized, an emergency or handbrakeapplication is effected by spring motor 325.

The second hydraulic motor 327 is normally pressurized through line 512by means of manifold 504. The first and second supply lines 511 and 512are interconnected by means of a pressure-responsive check valve 530mounted in manifold block 504. As a service brake application is made,line 160 is pressurized. The pressure at junction 518 not onlypressurizes line 511, but opens the pressure-responsive check valve 530to pressurize line 512. When valve 130 is open, hydraulic fluid underpressure flows to the second fluid motor 327 to displace spring motor325 to the left as illustrated in FIG. 7. When pressure is equalizedbetween lines 511 and 512, the pressure-responsive check valve 530closes. This prevents any further transfer of hydraulic fluid back fromline 512 to line 511 when the service brake application is terminated.After termination, the pressure in line 511 is essentially zero, whilethe pressure maintained in line 512 is equivalent to the full servicebrake application pressure. Since in actual practice, the service brakeapplication varies depending upon the weight of the load carried by thevehicle, the pressure present in lines 150 and 511 also vaires. Thepressure in line 512 however is equivalent to the highest service brakeapplication previously made, less any leakage from hydraulic motor 327and/or check valve 530. Since some leakage is inevitable, the systemrelies on each succeeding service brake application to restore thepressure in a second fluid motor 327 to a level suffi cient to preventthe application of spring motor 325.

The system also includes an exhaust valve for exhausting the fluidpressure maintained in conduit 512. When it is desired to make anemergency or handbrake application, exhaust valve 301 is opened and thepressure in conduit 512 is allowed to equalize with the pressure inconduit 511. It is important to note that the pressure present in thesecond hydraulic motor 527 is not dumped, but is rather equalized withany pressure present in the service brake line 511. If the pressure inconduit 512 were dumped, the emergency piston force from spring motor325 would add to the service brake force and result in a total brakeforce more than double that required to stop the car. This would simplylock the wheels and lengthen the stopping distance. That is, if aservice brake application is in effect at the time the emergency brakeapplication is made, the addition of the pressure generated by springmotor 326 to the existing service pressure would undoubtedly cause thebrake to lock and the wheels of the vehicle to slide.

It is essential in establishing the emergency brake force, that theminimum and maximum service brake pressures be accurately calculated. Itis necessary that the minimum service brake pressure that will besupplied under the lightest load conditions be sufficient to overcomethe spring motor pressure and restrain it under all conditions. It isalso necessary that the spring motor 325 be able to exert braking forceequal to that used in the service brake applications applied to the mostheavily laden cars. In other words, the spring motor must be capable ofsupplying a full service application under full load conditions, butmust be restrained from application by a pressure equivalent to thatapplied during the lightest of load conditions. This is done by usingthe differentially sized hydraulic motors 321 and 327. As illustrated inFIG. 2, the effective area of piston 322 is substantially less than theeffective area of piston 326. This enables the spring motor 325 to besized so that it will exert a pressure equivalent to the maximum serviceloading that will be present in motor 321. The smaller service pressureequivalent to the lightest load condition when presented to the largerservice area of piston 326 is still sufficient to compress spring motor325 and restrain it from application. If it is assumed that the fullservice brake application on the lightest car will be 600 pounds persquare inch, it would be desirable to size the second hydraulic motor327 to provide the full retraction of spring motor 325 when 550 poundsof pressure per. square inch is applied. The addition of any subsequentpressure up to and including 1,000 pounds per square inch will onlycause further compression of spring motor 325 toward its stop.

If the spring motor 325 is intended to provide a force equivalent to afull load service application, or a hydraulic pressure of 1,000 poundsper square inch, it will be necessary to reduce the fore applied by theemergency system if an emergency application is made during a servicebrake application. If, by virtue of the fact that the car is onlypartially loaded, the full service brake application pressure is 700pounds per square inch, the addition of a force equal to an extra 1,000pounds per square inch from spring motor would immediately lock thebrakes and cause the vehicle wheels to skid. To prevent this, theexhaust valve 301 equalizes pressure between the first and second supplylines 511 and 512.

If it is assumed that an emergency application is made during a serviceapplication, valve 301 is opened with approximately the same pressurepresent in both the first and second supply lines 511 and 512. Thepressure-responsive check valve 530 will insure that the pressure inline 512 is at least equivalent to the pressure in line 51 1. If anemergency application is made during a service brake application,equivalent pressures will be present in both lines, and the spring motor325 will be restrained from applying any additional pressure to piston322. This will prevent any excessive amount of service brake applicationto the brake pads 324 and 325. If however, the pressure is failing inthe service brake application line 511 and an emergency brake application is made, the spring motor will be applied when the 27 hydraulicpressure present in lines 511 and 512 drops below that normally requiredfor a full service application on a lightly loaded car. The amount ofbrake application will still be tempered, but will always produce abrake force greater than that required on an empty car.

If. for example, the pressure present in service line 511 drops to 300pounds per square inch, the differential between 300 pounds per squareinch and the application pressure for spring motor 325, that is 500pounds per square inch, will, when multiplied by the piston area of thesecond spring motor, give the force that the spring motor will produce.As the area of the second piston is roughly twice that of the servicepiston, pressure deficiency below the full service pressure will be madeup twice by the spring brake. Thus, in this example, the pressuredeficiency of 300 psi causes a spring brake application force equal toapproximately 500 psi to be added to the still existing 300 psi servicebrake pressure, thus producing braking force equivalent to a 900 psiservice pressure. The additional spring brake force will be applied topiston 322 by spring motor 325. Any further drop in the pressure presentin service line 511 will result in a further application of pressurefrom spring motor 525.

When it is desired to use the control system to actuate the handbrake,valve 301 is opened, and the hydraulic fluid in fluid motor 327 isallowed to dissipate into line 160. This insures that the parking brakewill remain fully applied even if the source of hydraulic pressure isde-energized, or if the pneumatic or hydraulic control means for thevehicle truck is disconnected. The mechanical retraction devicepreviously described with respect to FIG. 7 may be used to retract thehandbrake in the event it is desired to move the car without energizingthe brake control system.

The handbrake is released by actuating the service brake, and cyclingthe service brake through two or three applications. The cycling of theservice brake, and the imposition of fluid pressure on conduit 150 willpressurize fluid motor 327 through one-way valve 530 and the secondsupply line 512. In the preferred embodiment of the invention, the fluidmotor 327 is sized so as to fully release the spring motor 325 with asingle service application. However, it would be possible to use smallercapacity slack adjustors 41-43, and cycle the service brake system twoor three times to insure the hydraulic motor 327 is fully pressurized,and that spring motor 325 is fully retracted.

While I have thus described the preferred embodiments of the presentinvention, other variations will be suggested to those skilled in theart. It must be understood that the foregoing description is meant to beillustrative only, and not limitative of the present invention, and allsuch variations and modifications as are in accord with the principlesdescribed herein, are meant to fall within the scope of the appendedclaims.

I claim:

1. A brake system for vehicles having at least one axle with frictionand dynamic brakes and at least one axle with only friction brakes, saidsystem comprising: fluid operated friction brake means for said vehicle,said friction brake means including means responsive to variations influid pressure to actuate said friction brake means, said brake meanshaving:

first and second pressure-responsive fluid motors,

said first fluid motor being responsive to positive fluid pressurevariations to actuate said brake means;

spring motor means responsive to reductions in fluid pressures appliedto said second fluid motor to actuate said brake means,

a first fluid pressure control means for providing variations in thefluid pressure applied to the first pressure-responsive fluid motors ofsaid axle having friction and dynamic brakes, said first control meansresponding to a supplied signal which is inversely proportional to theamount of dynamic braking effort applied to said axle;

a second fluid pressure control means for providing fluid pressure tothe first pressure-responsive fluid motors of said axle having frictionbrakes, said second control means responding to a supplied signal whichis independent of the amount of braking effort applied to said axle withfriction and dynamic brakes; and

a third fluid pressure control means for providing reductions in fluidpressure to said second fluid motors, said third pressure control meansbeing responsive to both emergency braking control signals and parkingbrake control signals.

2. A brake system for vehicles as claimed in claim 1 wherein:

a. said first and second pressure-responsive fluid motors are hydraulicmotors, and

b. said first and second fluid pressure control means include apneumatic to hydraulic convertor for each pressure control means.

3. A brake system for vehicles as claimed in claim 2 wherein said firstand second control means comprises electro-pneumatic transducers fortranslating an electrical control signal to a pneumatic controlpressure.

4. A brake system as claimed in claim 1 wherein said first fluid controlmeans includes:

a. an electropneumatic transducer which responds to electricalvariations and control signals supplied by a controller for said dynamicbrakes, said transducer including a first and second torque motors, saidfirst torque motor responding to variations in said electrical controlsignal and said second torque motor responding inversely to variationsin the pneumatic output of said fluid control means, the combined outputof said torque motors producing variable plot pressures,

b. proportioning valve means for regulating the fluid pressure appliedby said transducer, said proportioning valve being responsive tovariations in said pilot pressures.

5. A brake system as claimed in claim 2 wherein said pneumatic tohydraulic convertor comprises a hydraulic motor means between said fluidpressure control means and said pressure-responsive hydraulic motors.

6. A brake system for vehicles as claimed in claim 1 wherein:

a. said first and second pressure-responsive fluid motors are hydraulicmotors, and

b. manifold means for directing fluid under pressure to said first andsecond fluid motors, said means including a first supply line for saidfirst fluid motor and a second supply line for said second fluid motor,

c. first valve means actuated by said third control means arrangedbetween and normally blocking communication between said first andsecond supply lines,

d. pressure-responsive valve means arranged between and establishingone-way communication 29 between said first and second supply lines,said valve means permitting a transfer of fluid under pressure from saidfirst supply line to said second supply line when the former exceeds thelatter.

7. An emergency and handbrake control system as claimed in claim 6wherein each of said motors defines a pressure-responsive working area,and the pressureresponsive working area for said second fluid motor islarger than the pressure-responsive working area for said first fluidmotor.

8. An emergency and handbrake control system as claimed in claim 6wherein said first valve means comprises an electro-magnetic valve whichestablished hydraulic communication between said first and second supplylines in response to an elelctrical control signal from said thirdcontrol means.

9. A brake system for vehicles as claimed in claim 6, wherein saidpressure-responsive valve means comprises a spring loaded ball checkvalve.

10. A brake system for vehicles as claimed in claim 1 wherein saidfriction brake means further includes a threaded retraction means forcompressing said spring motor means.

11. A brake system for vehicles as claimed in claim 10 wherein saidthreaded retraction means further includes:

a. means for manually rotating said threaded retraction means,

b. resilient means for rotating said threaded retraction means aftersaid spring motor has been retraeted,

c. means for resisting the rotation of said threaded retraction meansafter said spring motor has been retracted, said means providing for therelease and rotation of said threaded retraction means when said secondpressure-responsive fluid motor is actuated.

12. A brake system according to claim 2, wherein said pneumatic tohydraulic convertor comprises:

third pressure-responsive motor means comprising a housing defining avariable volume chamber having a moveable wall, said third motor meansbeing responsive to admission of pneumatic pressure to drive saidmoveable wall;

said housing also defining a fixed volume dump chamber adjacent saidvariable volume chamber for receiving a predetermined portion of theadmitted fluid pressure;

dump valve means responsive to an input control signal to opencommunication between said variable volume chamber and said dumpchamber; and

hydraulic cylinder and piston means operatively connected to saidmoveable wall for delivering hydrau- 30 lic fluid under pressure to saidfirst and second pressure-responsive fluid motors in response tomovement of said moveable wall. 13. A brake system according to claim12, wherein said dump valve includes means for simultaneously stoppingadmission of pneumatic pressure to said variable volume chamber whileopening communication between said variable volume chamber and dumpchamber.

14. A brake system according to claim 13, wherein said dump chamberincludes an orifice connecting it to atmosphere.

15. A brake system according to claim 5, further comprising a hydraulicslack adjuster operatively connected to said convertor, including areciprocating reaction element with first and second working chambers oneither side thereof, said first working chamber communicating with saidconvertor, said second working chamber communicating with said first andsecond pressure-responsive fluid motors; and transfer valve means forpermitting the transfer of hydraulic fluid from said first working spaceto said second working space when the volume in said second workingspace is at its minimum.

16. A brake system according to claim 1, wherein said first and secondpressure-responsive fluid motors comprise first and second pistons,respectively, further comprising:

two cooperating members connecting said first and second pistons, thefirst of said members being permitted substantial rotary movement andlimited axial movement, and the second of said members being permittedaxial but not rotary movement;

resilient means yieldably permitting greater axial movement of saidfirst member when said brake are applied by said spring motor than whensaid brakes are applied by said first fluid pressureresponsive motor;and

retraction means for retracting said handbrake piston, said retractionmeans including a threaded retraction screw rotatably, slidably mountedin said housing and threadably engaged for axial movement with saidsecond piston during parking brake application, said retraction screwcomprising abutting means for limiting axial movement of said screwduring retraction, whereby positive retraction of said second piston isprovided when said screw is rotated in a first direction, saidretraction means further comprising resilient drive means for rotatingsaid screw in the opposite direction when said secondpressure-responsive fluid motor is pressurized.

1. A brake system for vehicles having at least one axle with frictionand dynamic brakes and at least one axle with only friction brakes, saidsystem comprising: fluid operated friction brake means for said vehicle,said friction brake means including means responsive to variations influid pressure to actuate said friction brake means, said brake meanshaving: first and second pressure-responsive fluid motors, said firstfluid motor being responsive to positive fluid pressure variations toactuate said brake means; spring motor means responsive to reductions influid pressures applied to said second fluid motor to actuate said brakemeans; a first fluid pressure control means for providing variations inthe fluid pressure applied to the first pressure-responsive fluid motorsof said axle having friction and dynamic brakes, said first controlmeans responding to a supplied signal which is inversely proportional tothe amount of dynamic braking effort applied to said axle; a secondfluid pressure control means for providing fluid pressure to the firstpressure-responsive fluid motors of said axle having friction brakes,said second control means responding to a supplied signal which isindependent of the amount of braking effort applied to said axle withfriction and dynamic brakes; and a third fluid pressure control meansfor providing reductions in fluid pressure to said second fluid motors,said third pressure control means being responsive to both emergencybraking control signals and parking brake control signals.
 2. A brakesystem for vehicles as claimed in claim 1 wherein: a. said first andsecond pressure-responsive fluid motors are hydraulic motors, and b.said first and second fluid pressure control means include a pneumaticto hydraulic convertor for each pressure control means.
 3. A brakesystem for vehicles as claimed in claim 2 wherein said first and secondcontrol means comprises electro-pneumatic transducers for translating anelectrical control signal to a pneumatic control pressure.
 4. A brakesystem as claimed in claim 1 wherein said first fluid control meansincludes: a. an electropneumatic transducer which responds to electricalvariations and control signals supplied by a controller for said dynamicbrakes, said transducer including a first and second torque motors, saidfirst torque motor responding to variations in said electrical controlsignal and said second torque motor responding inversely to variationsin the pneumatic output of said fluid control means, the combined outputof said torque motors producing variable plot pressures, b.proportioning valve means for regulating the fluid pressure applied bysaid transducer, said proportioning valve being responsive to variationsin said pilot pressures.
 5. A brake system as claimed in claim 2 whereinsaid pneumatic to hydraulic convertor comprises a hydraulic motor meansbetween said fluid pressure control means and said pressure-responsivehydraulic motors.
 6. A brake system for vehicles as claimed in claim 1wherein: a. said first and second pressure-responsive fluid motors arehydraulic motors, and b. manifold means for directing fluid underpressure to said first and second fluid motors, said means including afirst supply line for said first fluid motor and a second supply linefor said second fluid motor, c. first valve means actuated by said thirdcontrol means arranged between and normally blocking communicationbetween said first and second supply lines, d. pressure-responsive valvemeans arranged between and establishing one-way communication betweensaid first and second supply lines, said valve means permitting atransfer of fluid under pressure from said first supply line to saidsecond supply line when the former exceeds the latter.
 7. An emergencyand handbrake control system as claimed in claim 6 wherein each of saidmotors defines a pressure-responsive working area, and thepressure-responsive working area for said second fluid motor is largerthan the pressure-responsive working area for said first fluid motor. 8.An emergency and handbrake control system as claimed in claim 6 whereinsaid first valve means comprises an electro-magnetic valve whichestablished hydraulic communication between said first and second supplylines in response to an elelctrical control signal from said thirdcontrol means.
 9. A brake system for vehicles as claimed in claim 6,wherein said pressure-responsive valve means comprises a spring loadedball check valve.
 10. A brake system for vehicles as claimed in claim 1wherein said friction brake means further includes a threaded retractionmeans for compressing said spring motor means.
 11. A brake system forvehicles as claimed in claim 10 wherein said threaded retraction meansfurther includes: a. means for manually rotating said threadedretraction means, b. resilient means for rotating said threadedretraction means after said spring motor has been retracted, c. meansfor resisting the rotation of said threaded retraction means after saidspring motor has been retracted, said means providing for the releaseand rotation of said threaded retraction means when said secondpressure-responsive fluid motor is actuated.
 12. A brake systemaccording to claim 2, wherein said pneumatic to hydraulic convertorcomprises: third pressure-responsive motor means comprising a housingdefining a variable volume chamber having a moveable wall, said thirdmotor means being responsive to admission of pneumatic pressure to drivesaid moveable wall; said housing also defining a fixed volume dumpchamber adjacent said variable volume chamber for receiving apredetermined portion of the admitted fluid pressure; dump valve meansresponsive to an input control signal to open communication between saidvariable volume chamber and said dump chamber; and hydraulic cylinderand piston means operatively connected to said moveable wall fordelivering hydraulic fluid under pressure to said first and secondpressure-responsive fluid motors in response to movement of saidmoveable wall.
 13. A brake system according to claim 12, wherein saiddump valve includes mEans for simultaneously stopping admission ofpneumatic pressure to said variable volume chamber while openingcommunication between said variable volume chamber and dump chamber. 14.A brake system according to claim 13, wherein said dump chamber includesan orifice connecting it to atmosphere.
 15. A brake system according toclaim 5, further comprising a hydraulic slack adjuster operativelyconnected to said convertor, including a reciprocating reaction elementwith first and second working chambers on either side thereof, saidfirst working chamber communicating with said convertor, said secondworking chamber communicating with said first and secondpressure-responsive fluid motors; and transfer valve means forpermitting the transfer of hydraulic fluid from said first working spaceto said second working space when the volume in said second workingspace is at its minimum.
 16. A brake system according to claim 1,wherein said first and second pressure-responsive fluid motors comprisefirst and second pistons, respectively, further comprising: twocooperating members connecting said first and second pistons, the firstof said members being permitted substantial rotary movement and limitedaxial movement, and the second of said members being permitted axial butnot rotary movement; resilient means yieldably permitting greater axialmovement of said first member when said brake are applied by said springmotor than when said brakes are applied by said first fluidpressure-responsive motor; and retraction means for retracting saidhandbrake piston, said retraction means including a threaded retractionscrew rotatably, slidably mounted in said housing and threadably engagedfor axial movement with said second piston during parking brakeapplication, said retraction screw comprising abutting means forlimiting axial movement of said screw during retraction, wherebypositive retraction of said second piston is provided when said screw isrotated in a first direction, said retraction means further comprisingresilient drive means for rotating said screw in the opposite directionwhen said second pressure-responsive fluid motor is pressurized.